Variable stroke dual plunger pump



J n 27, 196 c. F. GROM'ME I 3,327,632

VARIABLE STROKE DUAL PLUNGER PUMP Filed Jan. 4, 1966 6 Sheets-Sheet 1 INVENTOR. 10/. BY UA/eL [Geo/14,146,

f y $64 u ATTORNEYS.

June 27, 1967 c. F. GROMME 3,327,532

VARIABLE STROKE DUAL PLUNGER PUMP Filed Jan. 4, 1966 5 Sheets-Sheet 2 INVENTOR.

CA EL F GROMME,

A ATTORNEYS.

VARIABLE STROKE DUAL PLUNGER PUMP 0 V hi it Q [7 M v95 '9 44-"11 E1 :1 j Fig.8 44-)?) 3g v INVENTOR.

a 111g.16 BY CA/ZL FG/eoMMs, 44 "17 2 fqMW ATTORNEYS June 27, 1967 c. F. GROMME 3,327,632

VARIABLE STROKE DUAL PLUNGER PUMP Filed Jan. 4, 1966 5 Sheets-Sheet 4 INVENTOR. CA/ZL F GROMME,

m, fimi/m V ATTORNE-Ys United States Patent 3,327,632 VARIABLE STROKE DUAL PLUNGER PUMP Carl F. Gromme, Box 654, Kentfield, Calif. 94904 Filed Jan. 4, 1966, Ser. No. 518,587 14 Claims. (Cl. 103-2) This invention relates to a new and improved construction for a fluid pump, and relates more specifically to a fluid pump which is suitable as a distribution and injection system of liquid fuels for internal combustion engines whereby liquid fuel is delivered in controlled increments to such engines by either direct injections into the combustion chambers, injection into the intake manifolds at any desired location, or by direct injection into the intake valve chambers. The instant invention has particular utility in connection with fuel injection in conjunction with an internal combustion engine of the type taught in US. Letters Patent No. 3,071,123, although its utility is not so limited and the instant pump may be used with diverse types of internal combustion engines.

A principal object of the instant invention is to make possible a new variable stroke, sliding cam, dual plunger fluid pump for use in internal combustion engines to pump and distribute non-lubricating liquids in accurately metered and timed variable increments, in accordance with specific requirements.

Another object of the invention is to provide for an injection system for internal combustion engines which produces homogeneous mixtures of fuel and air during the normal intake processes of internal combustion engines under all conditions of operation.

Still a further object of the invention is to provide a fluid pump wherein a single compression stroke of one plunger may yield a plurality of discharges or pumping deliveries which create layered or stratified combustible charges within the cylinders of internal combustion engines.

Still another object of the instant invention is to achieve a fluid pump which is adaptable to internal combustion engines with single or multiple cylinders, either even or odd in number, and in which the distribution of the accurately metered increments of liquid pumped may be increased or decreased without major change in the size of the fluid pump, thus making it possible to use the fluid pump for different engine sizes.

Another object of the instant invention is to produce a fluid pump wherein the injection pressures may be chosen from ranges of low to high.

Still a further object of the invention is to provide for a fluid pump wherein rotating, sliding or rubbing parts operate at relatively slow speeds.

Yet a further object of the invention is to provide for a fluid pump wherein internal lubrication of moving parts may be accomplished by means of the engine lubrication system in the case of internal combustion engines.

Still a further object of the instant invention is to provide for a fluid pump which utilizes a balanced relief junction with a rotating distributor disk which has balanced fluid pressures.

Another object of the instant invention is to provide for a fluid pump wherein slidable cams having cam followers at an incline relative to the axes of the plungers are used to vary the stroke .of the plungers.

These and other objects of the invention, which will be described in greater detail hereinafter or which will be apparent to one skilled in the art upon reading the specifications, are accomplished by that certain construction and arrangement of parts of which the following describes an exemplary embodiment.

Reference is made to the drawings forming a part hereof and in which:

FIGURE 1 is a vertical longitudinal section through the entire device on line 1-1 of FIGURE 2.

FIGURE 2 is a representative cross section through the instant invention on line 2-2 of FIGURE 1.

FIGURE 3 is a sectional end view of one lobe of one cam taken on line 3-3 of FIGURE 1.

FIGURE 4 is a longitudinal section through one lobe of one cam taken on line 4-4 of FIGURE 3.

FIGURE 5 is a partial section taken on line 5-5 of FIGURE 1.

FIGURE 6 is a section taken on line 6-6 of FIG- URE 5.

FIGURE 7 is a plan section taken on line 7-7 of FIGURE 1.

FIGURE 8 is a representative section through a cam shaft support taken on line 8-8 of FIGURE 1.

FIGURE 9 is a plan view taken on parting line 9-9 of FIGURE 1.

FIGURE 10 is a reflected plan view taken along parting line 10-10 of FIGURE 1.

FIGURE 11 is a plan view of the top of the fluid pump of the instant invention showing a typical arrangement for an 8 cylinder engine.

FIGURE 12 is a section at enlarged scale through a plunger tip taken on line 12-12 of FIGURE 2.

FIGURE 13 is a section taken on line 13-13 of FIG- URE 12.

FIGURE 14 is a diagram at enlarged scale showing the relative positioning of distributor ducts, distributor ports and outlet ports for a fluid pump of the instant invention for a 4 cylinder engine.

FIGURE 15 is a diagram at enlarged scale showing the relative positioning of distributor ducts, distributor ports and outlet ports for a fluid pump of the instant invention for a 6 cylinder engine.

FIGURE 16 ia a diagram at enlarged scale showing the relative positioning of distributor ducts, distributor ports and outlet ports for a fluid pump of the instant invention for an 8 cylinder engine.

FIGURE 17 is a diagram at enlarged scale showing the relative positioning of distributor ducts, distributor ports and outlet ports for a fluid pump of the instant invention for a 5 cylinder engine.

FIGURE 18 is a detailed plan view at enlarged scale of outlet port inserts.

FIGURE 19 is a cross section taken on line 19-19 of FIGURE 18.

FIGURE 20 is a detailed section at enlarged scale taken on line 20-20 of FIGURE 11.

FIGURE 21 is a section through a typical injector valve.

FIGURE 22 is a partial plan section taken on line 7-7 of FIGURE 1 when a conservator is susbstituted in lieu of by-pass valves.

FIGURE 23 is a partial vertical section taken on line 23-23 of FIGURE 22.

FIGURE 24 is a section taken on line 24-24 of FIGURE 22.

3 FIGURES 25, 26, 27 and 28 are diagrams which illustrate the relationship of the injection periods to engine crankshaft revolution in 4, 5, 6 and 8 cylinder engines, respectively.

General organization and operation Reference is first made to FIGURESI and 2 wherein it may be seen that the fluid pump of the instant invention has a lower body 10 and an upper body 12 which are rigidly secured to each other along oil tight joint 14 by bolts or cap screws (not shown). Fluid to be pumped is delivered to the device from any suitable source of supply and is drawn through inlet 16 and dual check valves 17 above compression spaces 18 during the downward or suction strokes of alternately reciprocating plungers 20. Each valve outlet 22 of each compression space 18 is retained in its closed position during the downward or suction stroke of its respective plunger 20 by means of the suctional downward force created by the downward movement of the plunger.

During the compression or upward strokes of plungers 20 the fluid is ejected alternately through valve outlets 22 into ducts 24 in collector disk 26, the collector disk 26 being fitted with a sliding fiit in box 28 in the upper body 12 of the fluid pump. Ducts 24 lead to a centrally disposed circular recess on the upper surface of collector disk 26.

The fluid passes from collector disk 26 into a circular recess 30 in the under face of rotating distributor disk 32, which is axaially disposed over collector disk 26 and which rotates about pivot 34, and then through radial ducts 36 within the distributor disk 32. Radial ducts 36 terminate in ports 38in the upper face of the distributor disk 32.

Spur type gear teeth 40 are provided on the periphery of distributor 32 (integrally formed thereon if desired) and mesh with the teeth of drive gear 42.

Flow from ports 38 is distributed to and ejected through outlet ports 44 in cap 46 which is secured to the upper body 12 of the fluid pump by means of cap screws 48. Ports 44 feed directly into tubing connectors 50 which are connected by tubing. to injector or nozzle locations.

Plungers 20 are caused to reciprocate within sleeves 52 through interaction of cams 54, whichare subject to both sliding and rotational movement, and springs 56. Sleeves 52 are housed in the upper body 12 of the fluid pump and extend into collector disk 26, and as will be more fully explained hereinafter, cam followers 58 are angularly disposed with respect to earns 54 and their movement is guided by rigid guides 62, preferably of slotted, rectangular form, which are broached or otherwise cut in the lower body 10 of the fluid pump. To accommodate the sliding action of earns 54, rollers 60 of cam followers 58 have rounded peripheries. Springs 56 hold plunger spring supporting flanges 57, slotted members which fit snugly over grooves 59 in the plunger stems, in place.

Cams 54 slide upon shaft 64. Although shaft 64 is indicated as splined, a simple arrangement of key and keyway may also be used. Shaft 64 is relatively slender and is prevented from deflecting by cam support roller 65, supported by split collar 67, which rides upon race 69 formed on the bottom of lower body 10 (see FIGURES 1 and 8).

As is shown in FIGURES 1, 5 and 6, sliding motion is imparted to cams 54' by pivoted fork member 66 with rollers 68, each of which remain in contact with one of flanges 70' which are formed as an integral part of cams 54. Pivoted fork member 66 is rotated by means of lever 72 which may *be actuated by oil pressure, vacuum or manually, as, for example, by connection via linkage to the air throttle of spark ignition engines.

It should be noted that the horizontal component of the downward plunger force which tends to push earns 54 to the right (as seen in FIGURE 1) is countered by rollers 68 acting on flanges 70, and thus, consequently,

4 by lever 72. Also, compensation for altitude may be made through lever 72.

Rotational movement is imparted to cams 54 through an engine crankshaft or cam shaft which is operatively connected by means of an extension to shaft 64, partially illustrated by gear 74. Of course, it would be possible to accomplish a simplification of engine design by making only one unit with a mounting for both by mounting a spark distributor on the extension of shaft 64.

Rotation is imparted to distributor disk 32 by connection of shaft 64 to exemplary gear train 76, 78, and shaft 82.

Internal lubrication is accomplished by introducing a flow of lubricating oil from an external source, such as, for example, an engine lubricating oil supply, through lubricating; oil inlet fittting 84 to the reservoir formed by the lowerbody 10 of the fluid pump. Vertical connection passages 86 are provided to allow return of overflow lubricating oil to the source through overflow oil inlet fitting 87.

tion engines which use liquid fuels requires homogeneous mixtures during induction when the explosion cycle is used, regardless of whether the actual injection period is equal to or less than the full period of the piston motion during intake.

In compression ignition engines the air intake is not throttled and the requirements for meeting varyin speed and load are met by varying the fuel injections from minimum to maximum as a straight line function. In spark ignition engines the air intake is throttled, requiring varying fuel-air ratios for different conditions of speed and load. Spark ignition engines also generally require a richer mixture at idle, when residuals are high, than they do at higher speeds. However, at the highest speeds a richer mixture is desirable for maximum output.

With the foregoing considerations in mind, it will be seen from FIGURE 2 that earns 54 are formed as one sleeve-like unit which is slidable on shaft 64. Also, cams 54 are double lobed with each pair of oppositely disposed lobes being placed precisely at right angles to the other pair.

FIGURES 2, 3 and 4 show that the raised profile of cams 54 is such that the mixture of the fuel pumped after idle is gradually made leaner as the throttle is opened. To maintain the desired fuel-air ratios during ordinary operational speeds the cams 54 are profiled to yield a constant rate of increase of fuel. At higher speeds the cams are profiled to supply an increase of richness.

In order to attain continuity of fluid flow, the beginning of any compression stroke of one of plungers 20 must coincide precisely with the end of the compression stroke of the other plunger 20. Should it be desirable for plunger 20 to yield more than one injection per stroke, the entire plunger motion must be at uniform speed from start to finish, necessitating a uniform velocity cam.

Thepositioning of cams 54 on shaft 64 is of importance in increasing or decreasing the amount of fluid pumped with each plunger stroke. As shown in FIGURE 1, cams 54 are in contact with calm follower rollers 60 close to the mid-point of the total rise of one cam and at the base circle of the other cam. If cams 54 are moved to the right on shaft 64, the stroke of plungers 20 is decreased and the fluid pumped by each plunger stroke is diminished. Just the opposite result is effected if cams 54 are moved to the left.

In ordinary usage a sliding cam customarily operates with the cam follower axis normal to the cam axis. Such operation necessarily gives rise to :a unidirectional side thrust on the cam follower. Also, a customary sliding cam which is normal to the cam follower axis allows only line contact between the cam face and the cam follower roller, causing an uncertainty of action and a short life for the cam follower roller. The fluid pump of the instant invention avoids these difficulties because the cam follower axes are not placed at right angles to cams 54. This enables any side thrust that occurs to alternate in direction as cams 54 rotate or as they are varied in position along shaft 64. Also, the fact that the cam follower axes are placed at right angles to cams 54 allows the peripheral or circumferential faces of cam follower rollers 60 to be formed with a large radius of curvature, thereby permitting more than merely line contact between the surface of cams 54 and cam follower rollers.

FIGURE 4 shows a longitudinal profile of one of cams 54 at a point of one-half cam rise. The sequential lines 90 and 92 are conventionalized indications of points of contact between cam follower rollers 60 and the faces of cams 54 as cams 54 are shifted along shaft 64. The points of contact between cam follower rollers 60 and the rotating faces of cams 54 must follow compound curves, two of which are illustrated by dot-dash curves 94 and 96.

It should be noted that cam follower rollers 60 are virtually normal to cams 54 when cams 54 are in such a position that cam follower rollers 60 contact cams 54 along path 90. Also, when cam follower rollers 60 contact cams 54 on path 92, no reaction occurs except that force created by springs 56 tending to maintain contact between cams 54 and cam follower rollers 60.

Collector disk The upper ends of sleeves 52 in sockets 98 have sliding fits in the underside of collector disk 26. While it is important that leakage of pumped fluid around the tops of sleeves 52 be prevented, it is also necessary for the collector disk 26 to be free to move upwardly, maintaining constant contact with and pressure on distributor disk 32. This result is achieved by means of locked in place flanged seals 100 which fit snugly in bores 102 and by a plurality of springs 104 which are socketed in the underside of collector disk 26.

The plurality of springs 104 in the underside of collector disk 26 prevent leakage along interfaces 106 and 108 and tend to compensate for wear. The aforementioned seals 100 are preferably made of some flexible material which is impervious to the fluids handled. Additionally, it should also be pointed out that the constant pressures within the discharge systems act to distend seals 100 and further prevent leakage.

Difliculty will most likely be encountered in the placement of collector disk 26 during assembly of the instant device because of springs 184. T o alleviate this problem, threaded holes 110 are provided for set screws 112. The set screws 112 hold collector disk 26 in place after springs 104 are compressed. After assembly has been completed, set screws 112 must be loosened or replaced by a leak-off connection and tubing (not shown) in order to drain off any lubricating oil from around the collector disk.

Distributor disk and fluid distribution Fluid pressures exerted on distributor disk 32 are balanced by making the circular area of recess 30 in the distributor disk 32 equal to the sum of the area of ports 38. After the fluid pressures exerted on distributor disk 32 are balanced, the total pressure exerted by collector disk 26 on distributor disk 32 is equal to the summation of spring pressures due to springs 104 plus the relatively .slight upward pressure at seals 100. Springs 104 are selected to exert the pressure necessary to prevent separation of the face of collector disk 26 and distributor disk 32 under all conceivable operating fluid pressures.

, When injections are made through small constant area orifices a certain minimum pressure is required at idling speeds in order to achieve a reasonable flow of fiuid to an engine fuel injector. At high speeds, the minimum pressure required becomes very high inasmuch as such required pressure will vary inversely as the square of the orifice area and the time for injections. However, such high pressures are not required for fuel injection purposes in spark ignition engines and, therefore, a spring actuated pop-off injector may be utilized.

FIGURE 21 is an example of one type of spring actuated pop-off injector for use in manifold or valve chamber injection where the maximum injection pressure is virtually always the pressure required to open the injector. The spring actuated pop-off injectors 114 are preset for a predetermined injection pressure at idling. This insures an adequately small aperture through which to spray fuel into the air stream at idling speed, and also helps to largely determine the resulting pressures within the fluid pump from operation at all higher speeds.

Since spring actuated pop-off injectors are commonly provided with secondary filters, injector 114 is provided with pocket 116 in which a proper secondary filter may be placed. Additionally, it will be understood that in the cases of multiple cylinders, where there are exactly coincident injections as to start and finish, injectors 114 need not necessarily be matched as to opening pressure unless the injection periods are short enough to eliminate coincident injections, in which case the injectors must be matched in order to maintain identical injections of fuel through each injection.

In multiple cylinder engines when the injection periods are short enough to eliminate coincident injections, conditions are necessarily established where ports 44 and 38 are not in register, and the flow of the output of plungers 20 is blocked between injections. It is thus necessary to provide adjustable relief valves 118 which are interconnected to relief outlet 120 to allow passage of fluid back to the source. The spring tension of relief valves 118 is adjusted by threaded plugs 122. Since the opening pressure of relief valves 118 must be identical, plugged ducts 124 must be provided for installation of pressure gauges which are utilized during adjustment. Of course, in lieu of plugged ducts 124, matched springs held in place by cap screws may be used to hold valves 118 on their respective seats.

FIGURES 9, 10, 14, 15, 16 and 17 show a series of illustrative arrangements of ports 38 at the end of radial ducts 36 in distributor disk 32 in relation to outlet ports 44 in cap 46 for four, six, eight and five cylinder engines. It will be understood that fluid under pressure will be ejected only when ports 38 and 44 are in register during the traverse of ports 38 past ports 44. In FIGURES 9, 14, 15, 16, and 17, ports 38 are shown spaced 36 apart. This spacing is arrived at by dividing 360 by the number of ports 38. Other spacings, such as, for example, 360 divided by 5, '6, 7, 8, 9, etc., may be used depending upon the desired speed of rotation of the distributor disk 32, the relative speed ratio of distributor disk 32 to an engine crankshaft being equal to the reciprocal of the number of ports 38 for a two stroke engine and to the reciprocal of twice the number of ports 38 for a four stroke engine. Outlet ports 44 are spaced to permit ports 38 to traverse them in a proper even sequence.

FIGURES 25, 26, 27 and 28 illustrate in conventional diagram form the events in four stroke internal combustion engines for 4, 5, 6 and 8 cylinder units. These diagrams show the relationships of the injection periods, which are shown as extending for 180 degrees of crankshaft revolution and always beginning with the start of the intake strokes. The shaded rectangles represent the power strokes and the vertical, dashed group of lines represent the injection periods as coincident with the intake periods. The diagrams also show the overlapping effect when the injection periods last more than a certain number of degrees of crankshaft revolution. For example, it should be noted that overlapping injection periods occur in four stroke, four cylinder engines when the duration of injection periods corresponds to more than 180 of engine crankshaft rotation (360 divided by /z of the number of cylinders or 2). Likewise, overlapping injection periods occur in four stroke, six cylinder engines when the duration of injection periods corresponds to more than 120 of engine crankshaft rotation; in four stroke, eight cylinder engines when the duration of injection periods corresponds to more than 90 of engine crankshaft rotation; and in four stroke, five cylinder engines when the duration of injection periods corresponds to more than 144 of engine crankshaft rotation.

In cases of two stroke engines all injections will overlap. The extent of injection overlap is ascertained by the amount the piston stroke is occupied by the scavenging and air charging processes and the duration of injection during the balance of the compression stroke. It should be further understood that for an arrangement of an equal number of ports 38, two stroke engines require twice the speed of rotation of the distributor disk 32 as compared with four stroke engines.

The following explanation of one of the illustrative arrangements (FIGURE 14 which applies to a four stroke, four cylinder engine), is typical of all illustrated arrangements.

The direction of rotation of distributor disk 32 may, of course, be clockwise or counter-clockwise, and is governed by the rotation of earns 54 and the number of gears in gear train 76, 78' and 80 driving drive gear 42. The duration of injection periods corresponds to 180 of engine crankshaft rotation. The total distance of traverse of a port 38 past an outlet port 44 (the injection distance), must be (the speed ratio of distributor disk 32 to an engine crankshaft for a four stroke engine is equal tothe reciprocal of twice the number of ports, or 1 over 20) of 720 (the number of crankshaft revolutions for all cylinders to complete their cycles) divided by four (the number of cylinders), or 9". In the four stroke, four cylinder illustrative example of FIGURE 14, there will be no overlapping of injection periods when the duration of injection periods corresponds to 180 of engine crankshaft rotation. Therefore, one injection begins precisely as the last of four injections end.

All combinations of circumferential spacing of outlet ports 44 and ports 38, whether the duration of injection period is less than, equal to, or greater than 180 of engine crankshaft revolution, may be expressed by a rule,

which'is not affected by any overlapping of injection periods. Formulated the rule states that the circumferential spacing of adjacent outlet ports 44 is equal to that portion of the circumference of the circle which defines the locations of the ports 38 which is equal to the number of degrees subtended by the number of spaces between ports 38 corresponding to the number of engine cylinders less the circumferential width of one traverse of one port 38 past one port 44 the result being divided by the number equal to one less than the number of cylinders.

Using the four stroke four cylinder illustrative example of FIGURE 14 the circumferential spacing of adjacent outlet ports 44 may be computed in the following manner. The number of degrees subtended by the number of spaces between ports 38 corresponding to the number of engine cylinders is equal to 144. There are spaces between the 10 ports 38 each space comprising 36 degrees. FIGURE 14 is an illustrative arrangement for a four cylinder engine. Thus 144 is arrived at by multiplying 36 by 4. The circumferential distance of one traverse of one port 38 past one outlet port 44 was previously computed and found to be 9 degrees. The number equal to one less than the number of cylinders is 3. The circumferential spacing of ports 44 is thus equal to the quantity 144 minus 9 divided by 3 or 45.

The meshing of gear teeth on distributor disk 32 with the teeth of drive gear 42 need not follow any specific order. This is so because with a continuous fluid flow the identification of positions of plungers with any particular starting point of deliveries is not necessary.

It is of importance, however, that the positioning of distributor disk 32 with respect to cap 46 be precise and capable of being accurately checked and verified. To this end sight hole 126 (FIGURES 10, 11 and 20) is provided in cap 46 and witness slot 128 is provided in distributor disk 32. When sight hole 126 is centered over witness slot 128, pointer 130 rests directly on the center division of degree scale 131, indicating that a port 38 is in a position relative to a port 44 to just commence a traverse period. If the positioning of sight aperture 126 with witness slot 128 is accompanied simultaneously with a piston of an engine cylinder receiving a combustible charge, all subsequent injections will begin in proper sequences. Once positioning of sight apertures 126 with witness slots 128 has been satisfactorily accomplished, cap screw 132 is provided to seal the opening.

If it should be desired to advance or retard the beginning of the series of injections to each cylinder after the aforementioned positioning of sight aperture 126 and witness slot 128 has been accomplished, the cap screws 48 which project through slotted holes 134 in cap 46 may be loosened, allowing cap 46 to be rotated. If the rotation of distributor disk 32 is counter-clockwise, and the cap 46 is rotated clockwise or downwardly, as shown in FIGURE 11, moving the pointer 130 in a downward or clockwise direction, the beginning of a series of injections to each cylinder is advanced. Likewise, if the rotation of the distributor disk 32 is counter-clockwise and the cap 46 is rotated counter-clockwise or upwardly as shown in FIGURE 11, moving the pointer 130 in an upward or counter-clockwise direction, the beginning of a series of injections to each cylinder is retarded. Naturally, if the rotation of distributor disk 32 is in a clockwise direction, movement of cap 15 to advance or retard the series of injections to each cylinder will be just the opposite from that previously described.

Since the rotation of distributor disk 32 is at a relatively slow rate with respect to the engine crankshaft rotation, a one degree rotation of cap 46 with respect to the center mark at pointer 130 will correspond to a much greater actual degree of advancement or retardation. For example, for a four stroke engine a one degree rotation of cap 46 with respect to the center mark at pointer 130 results in an actual 20 degree adavncement or retardation.

Relation ofcam speed to distributor speed and to engine crankshaft speeds When it is desired that one stroke of one plunger 20 yield one injection only, the rotative speed of cams 54 relative to engine crankshaft speed is calculated in the following manner. For two stroke engines, it is the reciprocal of the number of cylinders divided by the number of cam lobes and for four stroke engines it is the reciprocal of the number of cylinders divided by one-half of the number of cam lobes. The device of the instant invention utilizes four cam lobes.

In the usual ranges of engine sizes the weight of an individual'fuel injection is very small, and consequently of little volume. It is thus desirable, as well as simple from a manufacturing point of view, to make the total quantity of fuel pumped during each plunger stroke a volume which may be conveniently handled, such as, for example, to provide for ten injections per plunger stroke. The speed ratio of cam rotation relative to crankshaft rotation when ten injections per plunger stroke are used, is one to ten or one to twenty for a four cylinder engine of two strokes or four strokes, respectively, which is the same as the exemplary relative speed of rotation of distributor disk 32 to the engine crankshaft. A six cylinder, eight cylinder, and five cylinder engine will require 15 injections, 20 injections and 12 /2 injections, respectively, to obtain the same speed ratio.

It should be noted that the use of a minimal acceleration period at the start of the cam action becomes acceptable at the slow actual rotational speeds indicated above. Another advantage of such slow rotational speeds is the reduction of impact and reduction of frictional heat and wear on all moving parts. Also, more time is permitted for operation of spring actuated check or relief valves, making it advantageous to provide a period of dwell on cams 54 between the end of the suction or intake strokes of plungers 20 and the beginning of their compression strokes, as shown by the segment of the cam base circle 55 in FIGURES 2 and 3.

The use of multiple injections per plunger stroke alternates the relative positions of plungers 20 during their compression strokes in relation to the sequence of the registering of ports 38 and 44. It necessarily follows that any variance in the quantities of separate increments of fuel passing through ports 38 and 44 which causes a variance in injected quantities to engine cylinders will occur in a pattern of succession, thus averaging out any unevenness in power output in the engine cylinders. This is particularly true when the number of injections per plunger stroke and the number of cylinders are not commensurate. It should also be noted that a delivery split between plunger strokes has no effect with regard to evenness of distribution when there are odd numbers of deliveries per plunger stroke, because the flow of fuel to the distributor disk 32 is always continuous.

At outlined above, even at high crankshaft revolutions, the movement of plungers 20, at least on the compression strokes, will be slow. Since the amount of leakage past a plunger is a function of the fluid density, pressure, clearance and time, sealing fits between sleeves 52 and plunger tips 136 are necessary. Also, since plunger compression spaces 18 are completely filled on the downward stroke of plunger 20, any slight movement of plungers 20 on a compression stroke will displace a certain corresponding amount of fluid. It therefore follows that care must be exerted in selecting the proper material for plunger tips 136 to avoid any appreciable degree of com pressibility which would affect deliveries.

It has been found that sealing fits between sleeves 52 and plunger tip 136 result if plunger tips 136 are made of a material which is relatively frictionless and slightly yielding, such as, for example, Teflon or nylon. These materials will operate successfully under pressures created during conditions of slight lubrication, such as is offered by gasoline and alcohol. However, under conditions of very high pressure, plunger tips 136 should preferably be made of an unyielding material and be lappedin in order not to add to volume changes due to fluid compressibility.

Plunger tips 136 are pinned or otherwise rigidly secured to plungers 20 and are formed with circular slots 138 within which are band springs 140 which maintain continuous close contact between the outer surface of plunger tips 136 and the bore of sleeves 52. This contact is, of course, reinforced by fluid pressure during the compressing stroke.

Conservator When the injection periods are short enough to cause intervals of no discharge; that is, when conditions are v established where ports 44 and 38 are not in register and the flow of the output of plungers 20 is blocked between injections, hydraulic lock will ensue, requiring the necessary relief afforded by adjustable relief by-pass valves 118 which are interconnected to outlet 120 from which any passed fluid is conducted back to the source. It should be noted, however, that the bypassing of pumped fluid represents a waste of performed work. Such bypassing of pumped fluid may be eliminated where there is no overlapping of injection period (for example, when the injection duration for four stroke, eight cylinder, six cylinder, or four cylinder engines is 90, 120, or 180 of engine crankshaft revolution, respectively, or, if for two stroke engines, /2 of the aforementioned injection periods) if a twin conservator (FIGURES 22 and 23) is substituted for relief by-pass valves 118.

When a twin conservator is substituted for relief valves 118, resultant blocked flow between injections is forced through ducts 142 into chambers 144 wherefluid pressure builds up instantaneously. The fluid pressure in chamber 144 depresses fluid-tight diaphragm 146, overcoming the resistance of spring 148 against head 150 of spindle 152.

The fluid pressure required to compress spring 148 must be greater than the pressure required to open the springactuated pop-off injector 114 so that when ports 38 begin to traverse ports 44, accumulated fluid will be ejected simultaneously with the balance of the injection displacement of plungers 20. This necessarily results in no fluid by-pass and full utilization of the displacement of plungers 20.

Even though the revolution of distributor disk 32 and the reciprocation of plungers 20 may be relatively slow, the back and forth movement of fluid-tight diaphragms 146 will occur once for each injection. At high engine speeds, such as, for example, 6,000 r.p.rn. for an eight cylinder, four stroke engine, the back and forth movement of fluid-tight diaphragms 146 will occur once each .0025 second. It should be noted, however, that since the area of fluid-tight diaphragm 146 may be made larger, the actual movement thereof can be diminished so that it will be in the nature of a flutter.

In most instances it is preferable to control the maxi mum extent of the movementof fluid-tight diaphragms 146. This is accomplished by providing a positive stop for the return action of fluid-tight diaphragms 146 and spindles 152 which move within drilled, threaded plugs 154 under action of springs 148. The positive stop effect results from drilled, threaded plug 154, a stop nut 156 on threaded end 158 of spindle 152, all housed in sleeves 160 which are held in place by hold-down bolts 164. The tension on springs 148 may be adjusted by screwing drilled, threaded plugs 150 within sleeves 160. At the same time, stop nut 156 is screwed on the threaded end 158 of spindle 152 toward head 150 so that head 150 is just short of touching the fluid-tight diaphragm 146.

The rapid movement of fluid-tight diaphragms 146 and spindle 152 requires lubrication. This may be accomplished by partially filling chamber 166 with suitable lubricating oil through threaded aperture 168 in cap 162 when thumb nut 170 is removed.

Adaptability of this invention It is important that limitations of application of the instant invention be minimized as much as possible. To this extent it has been found that oneset of cams within one size body and one size cap may be utilized with engines of different sizes and capacities. Obviously, when the plunger strokes are fixed by cam size the displacement volumes of plungers 20 may only be changed by increasing or decreasing the bores of sleeves 52. This is accomplished by utilizing sets of matched sleeves 52 and corresponding sets of plungers 20 and plunger tips 136. r A further means of increasing the adaptability of the instant invention is the provision of inserts 172 or ring shaped insert 174. The inserts 172 or ring shaped insert 174 are provided with ports 176 corresponding to cut in place ports 44. The purpose of inserts 172 or ring shaped insert 174 is to provide a simple means of varying the sizes and lengths of ports 44 without changing cap 15. For example, when a shorter injection period is desired, inserts with ports of shorter circumferential lengths may be installed. While shown in FIGURE 10 for an eight cylinder engine, it will be understood that inserts 172 or ring shaped insert 174 may be provided for other numbers of discharge ports 44.

It will be understood that modifications may be made without departing from the spirit of the invention and therefore no limitations other'than those specifically set forth in the claims are intended or should be implied.

The embodiments of the invention in which an exclu- I I sive property or privilege is claimed are defined as follows:

1. A fluid injection pump for use in internal combustion engines, which comprises:

(a) a hollow pump body;

(b) a drive shaft within said hollow pump body;

(c) means for imparting rotational movement to said drive shaft;

(d) a pair of integrally connected double lobed cams slidably mounted on said drive shaft, said cams being operatively connected to said drive shaft for rotation therewith, said cams being disposed normal with respect to each other, the surfaces of said cam lobes being angularly disposed with respect to the longitudinal axis of said drive shaft;

(e) means for imparting sliding movement to said cams;

(f) a pair of sleeves within said hollow body which form two compression spaces;

(g) a pair of plungers slidably mounted within said sleeves, said sleeves and plungers having the longitudinal axes inclined with respect to the longitudinal axis of said drive shaft;

(h) cam followers at the ends of said plungers in contact with the surfaces of said cam lobes;

(i) means for supplying fluid from a source of supply to said compression spaces;

(j) a collector disk overlying and in communication with said sleeves, said collector :disk having two ducts, one from each of said sleeves, which lead to a central recessed area in its upper face;

(k) a rotatable distributor disk juxtaposed to said collector disk, said distributor disk having a recessed area in its under face which mates with the central recessed area in the upper face of said collector disk and a plurality of ducts projecting radially from the recessed area in said distributor disk, each of which terminates in a port in the upper face of said distributor disk;

(1) gear train means operatively connecting said dis tributor disk to said drive shaft to impart rotational movement to said distributor disk;

(m) a cap atop said hollow body overlying said distributor disk, said cap having a plurality of outlet ports corresponding to the number of cylinders of the internal engine utilizing said fluid pump;

(n) means for aligning the ports in the distributor disk relative to the outlet ports in said cap; and

() overflow means in communication with said compression spaces to provide for fluid which has not been discharged from said compression spaces during each stroke of said plungers.

2. The fluid pump according to claim 1 wherein said overflow means comprises a pair of spring adjustable relief valves'which are interconnected to an outlet leading to the fluid source.

3. The fluid pump according to claim 1 wherein the surface area of the central recess in the underface of said distributor disk is substantially equal to the sum of the surface areas of the ports in the upper face of said distributor disk so that the fluid pressures are balanced.

4. The fluid pump according to claim 1 wherein the means for aligning the ports in said distributor disk with the outlet ports in said cap comprises an aperture in said cap in communication with said distributor disk, said distributor disk having a witness slot therein, and a pointer upon said cap with a scale division thereunder so that when said aperture is centered over the witness slot the pointer rests directly upon a specific division of the scale and a leading edge of a port in said distributor disk is in alignment with the leading edge of an outlet port in said cap..

5. The fluid pump according to claim 1 wherein said overflow means comprises twin conservators, each of which comprises a chamber having a duct leading to it from said compression space, a fluid-tight diaphragm at the end of said chamber, and means for regulating the yield of said fluid-tight diaphragm so that when the ports in said distributor disk begin to traverse the outlet ports in said cap the accumulated fluid is ejected simultaneously with the balance of the fluid displaced by said plungers.

6. The fluid pump according to claim 1 wherein a separate insert is fitted into each outlet port in said cap so that the surface area of the outlet port may be varied in order to obtain different injection periods.

7. The fluidpump according to claim 1 wherein a single ring-shaped insert with ports corresponding to the outlet ports in said cap is fitted over the outlet ports of said cap so that the surface ofsaid outlet ports may be varied to obtain different injection periods, the ports in said insert having surface areas smaller than the surface areas of said outlet ports.

8. The fluid pump according to claim 1 wherein the means for imparting axial movement to said cams comprises a pair of spaced apart flanges formed as an integral part of one of said cams; a pivoted fork member, the furcations of the forked member engaging said last named cams between said flanges, and means pivotally mounting said forked member so that said forked member will act against said flanges and cause said cam to move axially along said drive shaft.

9. -A fluid injection pump for use with internal combustion engines which comprises:

(a) a hollow pump body;

(b) a drive shaft rotatably mounted within said pump body;

(c) means for imparting rotational movement to said drive shaft;

((1) a pair of bores in said pump body;

(e) a pair of plungers mounted within said bores for reciprocal movement relative thereto, the upper ends of said plungers defining compression spaces at the ends of said bores, said bores and plungers having their longitudinal axis angularly disposed with re spect to the longitudinal axis of said drive shaft;

(f) cam followers operatively connected to the opposite ends of said plungers;

(g) a pair of double lobed cams slidably mounted on said drive shaft and operatively connected to said shaft for rotation therewith, said cams being positioned to be contacted by said cam followers and having their lobes oppositely disposed with the lobes of one cam disposed normal to the lobes of the other, said lobes having inclined cam surfaces extending in the direction of the length of said drive shaft, whereby upon axial movement of said cams relative to said drive shaft, the stroke of the plungers will be varied;

(h) means for supplying fluid from a source of supply to the compression spaces in said bores;

(i) discharge ducts at the remote ends of said compression spaces for discharging fluid therefrom, whereby upon reciprocating movement of said plungers fluid will be drawn into said compression spaces and discharged into said discharge ducts;

(j) a collector disk in communication with said discharge ducts for collecting fluid discharged into said ducts, said collector disk having an outlet orifice in the opposite face thereof;

(k) a rotatable distributor disk juxtaposed to the said opposite face of the collector disk, said distributor disk having an inlet orifice in its undersurface in communication with the outlet orifice in said collector disk and a plurality of ducts connected at their inner ends to said inlet orifice and terminating at their opposite ends in a plurality of ports in the outer surface of said distributor disk;

(1) drive means operatively connecting said distributor disk to said drive shaft for rotation therewith; and

(m) a cap overlying said distributor disk, said cap having a plurality of discharge ports therein positioned to be sequentially brought into communication with the ports in said distributor disk as said disk is rotated, said discharge port-s being adapted to be connected to the cylinders of an internal combustion engine.

10. The fluid pump claimed in claim 9 wherein said collector disk is spring-biased into contact with said distributor disk, whereby said collector disk maintains constant contact with and pressure on said distributor disk.

11. The fluid pump claimed in claim 9 including overflow means in communication with the compression spaces in said bores to provide for fluid which has not been discharged from said compression spaces during each stroke of said plungers.

12. The fluid pump claimed in claim 9 wherein the surface area of the outlet orifice in said collector disk is equal to the surface area of the inlet orifice in said distributor disk, and wherein the sum of the surface areas of the ports in said distributor disk is equal to the surface area of its said inlet orifice.

13. The fluid pump claimed in claim 9 wherein the means for supplying fluid from a source of supply to the compression spaces in said bores includes a check valve oriented to prevent back flow therethrough, and wherein normally closed check valves interconnect the discharge ducts at the ends of said compression spaces and said collector disk.

14. The fluid pump claimed in claim 9 wherein the means for imparting aX-ial movement to said cams comprises a pair of spaced apart flanges operatively connected to said cams, a pivoted fork member mounted within said housing, the furcations of said forked member engaging between said spaced apart flanges, and means pivotally mounting said forked member so that said forked memher will act against said flanges and thereby effect axial displacement of said cams relative to said drive shaft.

References Cited UNITED STATES PATENTS 2,684,630 7/ 1954 Widmer et a1 103-2 2,851,953 9/1958 Peterson 103-2.1 3,016,837 1/1962 Dlugos 103-37 3,064,579 11/1962 Staege et al 1032.1 3,065,699 11/ 1962 Gromme 103-21 3,150,595 9/1964 Bernard 103-37 DONLEY J. STOCKING, Primary Examiner. W. J. KRAUSS, Assistant Examiner. 

1. A FLUID INJECTION PUMP FOR USE IN INTERNAL COMBUSTION ENGINES, WHICH COMPRISES: (A) A HOLLOW PUMP BODY; (B) A DRIVE SHAFT WITHIN SAID HOLLOW PUMP BODY; (C) MEANS FOR IMPARTING ROTATIONAL MOVEMENT TO SAID DRIVE SHAFT; (D) A PAIR OF INTEGRALLY CONNECTED DOUBLE LOBED CAMS SLIDABLY MOUNTED ON SAID DRIVE SHAFT, SAID CAMS BEING OPERATIVELY CONNECTED TO SAID DRIVE SHAFT FOR ROTATION THEREWITH, SAID CAMS BEING DISPOSED NORMAL WITH RESPECT TO EACH OTHER, THE SURFACES OF SAID CAM LOBES BEING ANGULARLY DISPOSED WITH RESPECT TO THE LONGITUDINAL AXIS OF SAID DRIVE SHAFT; (E) MEANS FOR IMPARTING SLIDING MOVEMENT TO SAID CAMS; (F) A PAIR OF SLEEVES WITHIN SAID HOLLOW BODY WHICH FORM TWO COMPRESSION SPACES; (G) A PAIR OF PLUNGERS SLIDABLY MOUNTED WITHIN SAID SLEEVES, SAID SLEEVES AND PLUNGERS HAVING THE LONGITUDINAL AXES INCLINED WITH RESPECT TO THE LONGITUDINAL AXIS OF SAID DRIVE SHAFT; (H) CAM FOLLOWERS AT THE ENDS OF SAID PLUNGERS IN CONTACT WITH THE SURFACES OF SAID CAM LOBES; (I) MEANS FOR SUPPLYING FLUID FROM A SOURCE OF SUPPLY TO SAID COMPRESSION SPACES; (J) A COLLECTOR DISK OVERLYING AND IN COMMUNICATION WITH SAID SLEEVES, SAID COLLECTOR DISK HAVING TWO DUCTS, ONE FROM EACH OF SAID SLEEVES, WHICH LEAD TO A CENTRAL RECESSED AREA IN ITS UPPER FACE; (K) A ROTATABLY DISTRIBUTOR DISK JUXTAPOSED TO SAID COLLECTOR DISK, SAID DISTRIBUTOR DISK HAVING A RECESSED AREA IN ITS UNDER FACE WHICH MATES WITH THE CENTRAL RECESSED AREA IN THE UPPER FACE OF SAID COLLECTOR DISK AND A PLURALITY OF DUCTS PROJECTING RADIALLY FROM THE RECESSED AREA IN SAID DISTRIBUTOR DISK, EACH OF WHICH TERMINATES IN A PORT IN THE UPPER FACE OF SAID DISTRIBUTOR DISK; (L) GEAR TRAIN MEANS OPERATIVELY CONNECTING SAID DISTRIBUTOR DISK TO SAID DRIVE SHAFT TO IMPART ROTATIONAL MOVEMENT TO SAID DISTRIBUTOR DISK; (M) A CAP ATOP SAID HOLLOW BODY OVERLYING SAID DISTRIBUTOR DISK, SAID CAP HAVING A PLURALITY OF OUTLET PORTS CORRESPONDING TO THE NUMBER OF CYLINDERS OF THE INTERNAL ENGINE UTILIZING SAID FLUID PUMP; (N) MEANS FOR ALIGNING THE PORTS IN THE DISTRIBUTOR DISK RELATIVE TO THE OUTLET PORTS IN SAID CAP; AND (O) OVERFLOW MEANS IN COMMUNICATION WITH SAID COMPRESSION SPACES TO PROVIDE FOR FLUID WHICH HAS NOT BEEN DISCHARGED FROM SAID COMPRESSION SPACES DURING EACH STROKE OF SAID PLUNGERS. 